Multi-speed transmission with two planetary gears

ABSTRACT

The invention relates to a multi-speed transmission on a central axis (6a), comprising a drive-side hollow shaft (4a) and an output-side hub sleeve (I) and two, at least two-stage, planetary gearboxes (EG, NSG), arranged coaxially therebetween or therein, each gearbox having sun gears, planetary gears which are connected to one another via associated frames, and ring gears, and each can be shifted into a direct drive and into a plurality of stepped conversion modes, wherein the first planetary gearbox (EG) is driven on its frame (8) and can be selectively shifted, by means of three clutches (20, 30, 40), into the direct drive or by axially fixing individual sun gears (9c,9d) into said transmission modes, and the second planetary gearbox (NSG) can be shifted by further clutches (50, 60, 70, 80, 90) by means of an axial fixing of the sun gear (204) when the input drive and output drive are shifted between its ring gear (200) or frame (202) into a conversion mode at a low speed and into a conversion mode at high speed and directly into a gear. By means of an axially displaceable control slide, spring-loaded, radially extending coupling rings (e.g 91a, 91b) of a respective gear stage can be released, such that the control slide is guided in helical circular grooves (24 . . . 94) of a coaxial switching drum (6b) and in radial slots that are oriented parallel to the gear axis of a hollow axle (6a) enclosing them, respectively actuating the associated coupling rings.

In accordance with the generic term of the main claim, the inventionrelates to a multi-speed transmission, in particular for a two-wheeler,which is mounted on a central axle, has a hollow shaft on the input sideand a hub sleeve on the output side, between which two at leasttwo-stage planetary gears are inserted, which are each connected to oneanother by sun gears and associated webs, and each of which can beshifted with a ring gear either into a block revolution or one ofseveral transmission or reduction modes.

A similar multi-speed transmission is known from EP 0915 800 B1. Thiscomprises an axle which can be mounted non-rotatably on a bicycle frame,a driver rotatably mounted on the axle, a sleeve rotatably mounted onthe axle, a gear shift mechanism provided with a first and a secondplanetary gearbox arranged in the sleeve and coupled to the driver andthe sleeve to provide multiple gear ratios between the driver and thesleeve, the planetary gearboxes comprising at least two sun gearscoaxial with the axle, have at least one planet carrier withcorrespondingly stepped planet gears engaging with the two sun gears andat least one ring gear engaging with the planet gears as well as ashifting device for selecting the transmission ratios by selectivelyconnecting the sun gears to the axle, for which purpose the shiftingdevice has means for selectively connecting the ring gear or the planetcarrier of the first and/or second planetary gearbox to a sun gear ofthe same planetary gearbox. This preferably results in a 7×2=14-speedgearbox, in which a seven-speed gearbox, consisting of a direct gear andtwo two-stage manual gearboxes, is coupled in series with anothergearbox, which has a direct gear and a shiftable planetary arrangement.The gears are shifted with 4 pawls, 3 axial clutches and 6 planetarysets, which are arranged to form three two-stage planetary arrangements.The disadvantage of this arrangement is the low input torque, whichresults in a preferred chain or belt gear ratio of preferably 2.5. Thesun gears are axially fixed with pawls, which requires an internalgearing of the same, which is loaded asymmetrically by the pawls—forspace reasons only one per sun gear—and only on one internal tooth. Afurther disadvantage is that no thru axle can be used.

A further multi-speed transmission is known from EP 2 028 096 A1, inwhich only 5 planetary sets, formed by a double and a triple steppedplanet, but 12 clutches are provided.

The disadvantage of this arrangement is that six of the clutches arepawls which serve to fix the axes of the 4 sun gears. Other clutches,which act axially, have approximately ring gear diameters. In addition,the ring gears must be supported on their large diameters. The shiftingmechanism of the transmission is very complex and costly to manufacture.In addition, the stepped planets cannot be supported by roller bearings,since some planets have a small number of teeth (14 or 15 teeth). A 12mm thru axle is also not realizable. The input gear, consisting of twosun gears, two hollow gears and three triple stepped planets, providesthree fast gear ratios in addition to the direct gear. The first gear ofthis transmission is the direct gear, which is formed by the coupling ofplanet carrier and ring gear. In the second and third gear, thetransmission forms two sub-transmissions, but these are different foreach of the two gears. The fourth gear is a simple planetary gearboxdriven by the planet carrier with a double-step planet. The overalldesign results in a high number of different complex parts and a totalmass that is significantly higher than the mass of the gearbox knownfrom EP 0 915 800 B1. Furthermore, multi-speed hub gears with three orfour planetary sets are known from U.S. Pat. No. 9,279,480 B2, which canbe coupled in various ways by nine clutches.

The clutches are actuated by a camshaft which actuates radial shiftingmeans.

The disadvantages here are the low input torque that can be toleratedand the large number of different, complex production parts.

In the state of the art, transmissions of this type are shifted withpawls, which are supported in apertures in a transmission axle,controlled by cams on a shift shaft, and folded out or in (cf. EP 2028096 A1; DE 10 2010 051.727 A1). If a gear wheel is to be fixed to theaxle, the pawl folds out and engages in an internal toothing of the gearwheel to be fixed. On the one hand, the internal gearing has a notcheffect, and on the other hand, the gear wheel is subjected to punctualinternal loading. The wall of the hollow axle must be relativelythick-walled due to the high forces and thus have a relatively largediameter. It is known that pawls are difficult to collapse under load. Afurther switching type is shown in the U.S. Pat. No. 9,279,480 B2, whereradial cams actuate the switching means.

Common to all these well-known hubs is that the drive pinion mounted ona hollow shaft has about half—or considerably less—the number of teeththan a pedal crank output blade usually has, because the input forcesthat can be tolerated are relatively low.

The toothed belts that are now widely used as a means of transmittingpower to the rear wheel require a significantly larger number of teethon the rear wheel because of their rigidity. The e-bikes that have beenon the market for several years now enable higher average speeds, whichin turn require a high transmission ratio bandwidth. Meanwhile, the 5 mmquick-release axle, which is used to clamp a hub in the rear triangle orfork, is increasingly being replaced by thru axles. In the rear wheel adiameter of 12 mm is common. The thru axle is mostly a long screw with ashort 12 mm thread. The thru axle is inserted from one side through asuitable hole, passed through the hollow hub axle and then screwed intothe thread of the opposite drop out. When tightening, the connection ispreloaded.

The invention is based on the task of designing the multi-speedtransmission in such a way that a wide range of transmission ratios canbe achieved with low weight, high efficiency, simple manufacture, atransmission ratio suitable for toothed belts and under load, withsimple control, particularly good shifting and can be mounted in thedropouts of the rear end by means of a thru axle.

The solution is indicated in the identification of the main claim.Advantageous designs are indicated in the sub and secondary claims.

The invention is explained in the following with reference to FIG. 1a/b, FIG. 2 a/b, using an example of design.

A preferred design is a nine-gear transmission with constant gear stepsof approx. 24% at a transmission ratio range of approx. 560%. The inputtransmission ratio is selected in such a way that with the secondarytransmission ratio of the chain or belt drive with e.g. 40 teeth at thefront and a rider with e.g. approx. 36 teeth at the transmission hub, atotal transmission ratio i_ges for the first gear of approx. i_ges.=1.7results. For the ninth gear this results in a transmission ratio of0.30.

This is achieved by coupling two three-speed planetary gears.

A first transmission is happy to exist. FIG. 1a consists of a web 8 withdouble stepped planets 9 a and 9 b, a ring gear 7 a with side wall 7 band two sun gears 9 c and 9 d, which mesh with the correspondingplanets. The bridge 8 is driven by a driver 4 a, on which the rittling 2is fixed rotationally. The output of the first gear is via the ring gear7 a. If the transmission is operated in block rotation, the direct gearwith i_1=1 is obtained, and the gear ratio with i_2=1/1.24=0.8065 isobtained if the smaller of the two sun gears is fixed to the axis. Thegear ratio with i_3=1/1.5376=0.65 is obtained when the larger of the twosun gears is fixed to the axis. Thus, the transmission ratio is fast.

A first preferred design of a second gear consists of two identicalplanetary gearboxes, mirrored to each other, each consisting of a ringgear, a sun gear and double stepped planets, which are coupled by meansof bridges that are connected to each other in a rotationally fixedmanner (not shown). The first subgear of this add-on set is a reductiongear with approx. i_4=10.24{circumflex over ( )}3=1.91 and the secondsubgear is a transmission gear with i_6−I/i_4=0.5245. The direct gearwith i_5=1 is achieved when both subgears rotate as a block.

A second preferred version of a second transmission according to FIG. 2aconsists of a planetary gearbox identical to a partial gearbox of thefirst preferred version. The three gears of the first preferred designare realized by switching the gear input between ring gear and web. Thesame procedure is used for the abrasion. This does not change the gearratios.

The following table shows that in the case of gears G1, G5, G6, G7, onlyone planetary gear set is involved in the power transmission and in thecase of gears G2, G3, G8 and G9 only their two. This stands for a goodefficiency of the gear hub. Gear G4 does not require any rolling motionof the gears.

Transmission Transmission EG NSG Gear i_1 i_2 i_3 i_4 i_5 i_6 i_Gn 1 11.91 i.si 2 0.81 1.91 1.55 3 0.65 1.91 1.24 4 1 1 1 5 0.81 1 0.81 6 0.651 0.65 7 1 0.5245 0.53 8 0.81 0.5245 0.43 9 0.65 0.5245 0.34

According to the invention, the multi-speed transmission is shifted withaxial clutches. A clutch is formed by two opposing discs, which can beoperated by means of intermediate elements, with axially acting ratchetteeth acting in one direction, mounted on a circular ring perpendicularto the axis of rotation. In the following, a coupling that triggers theblock rotation, “block lock”, which prevents the sun gear from rotatingwhen driven by the web, “tracking lock” and which prevents the sun gearfrom rotating backwards when driven by the ring gear, “backstop”. Inthis case, the same direction of rotation as the driving belt orsprocket wheel is meant. If torque is to be transmitted with the samedirection of rotation, the term “driver” is used.

In the following table the symbols used in the figures are listed withtheir meaning. In FIGS. 1b and 2b , the reference symbols of thecouplings are counted through in steps of ten and those of theassociated parts or elements in steps of one. In FIGS. 4, 5 and 6, whichdescribe the shifting states of the gears G1, G5, G9, only the clutchnumbers are then mentioned.

Sliding ring with cylindrical pin and outer collar, axially movable,rotation on a shaft or axle blocked by internal gear teeth

Slide ring with inner collar, axially movable, rotation in a shaft oraxle blocked by external toothing.

Slotted hole in a hollow axle or hollow shaft to allow a cylindrical pinto pass through; represented by missing hatching and outward pointingsquare brackets.

Circumferential groove in the control shaft for guiding and axialmovement of the cylindrical pins a)

 b)

Backstop or driving clutch, coupled in Direction of rotation of thedrive a) open, b) closed n)

 b)

Travel stop, double against the direction of rotation of the drive a)open, b) closed a)

 b)

Rolling bearing single, double

Locks, axially displaceable with cylindrical pin, rotationally fixed byinternal toothing

Locks, axially fixed by circlips, rotationally fixed by internal gear

detachable connection of two components a)

 b)

bidirectionally acting clutch; a) unswitched, b) switched

The preferred rear wheel hub, FIG. 1a,b and FIG. 2a,b , consists of ahub sleeve 1 supported by roller bearing 3 a on the drive 4 a and byroller bearing 3 b on axle 6 a. A drive pinion 2 for belt or chain driveis located on the drive 4 a, rotationally fixed but detachably connectedto it. The roller bearing 5 supports the drive 4 a on the axle 6 a.Inside the axle 6 a there is a switching drum 6 b. The oblong holes 101,102, 103 and 104 are arranged parallel to the rotation axis and allowthe cylindrical pins acting as actuators to pass through into thecircumferential grooves 24, 34, 44, 54, 64, 74, 84 and 94 of theswitching drum 6 b. The grooves are designed in such a way that theintended movements of the shifting pins for shifting the individualgears are executed. The compression springs 22, 32, 42, 52, 62, 72, 82and 92 are arranged in such a way that the clutches are always openedagainst the force of these springs. In an advantageous variant, acomplete control structure is arranged on 180 degrees of thecircumference of the shift drum and the hollow axle. This results in thegreat advantage that an exact copy of the same can be arranged on thesecond half of the circumference and thus the clutch plates can becontrolled in parallel by means of second cylindrical pins. Thisprevents any shift inhibition due to a tilting load, which inevitablyoccurs with one-sided actuation.

The second NSG gear unit, FIG. 2a , consists of a drive sleeve 4 b withinternal and external toothing and at least one, but preferably two orthree oblong holes 4 c offset by 180 degrees or 120 degrees, an outputsleeve 4 d also with internal and external toothing and at least one,but preferably two or three oblong holes 4 e offset by 180 degrees or120 degrees. The drive sleeve is connected to the side wall 7 b via theexternal gearing, and the abrasion sleeve is connected to the hub sleeve1. The gear stage itself consists of the ring gear 200, the side walls201 a and 201 b, at least one of which is designed to be connected tothe ring gear 200 via a detachable connection, and also of the web 202and the planets 203 a and 203 b, which are connected to each other in arotationally fixed manner, and the sun gear 204. The planet 203 a mesheswith the ring gear 200 and the planet 203 b with the sun gear 204, theside wall 201 a is connected with the coupling part 51 b and the sidewall 201 b with the coupling part 91 b in a rotationally fixed manner.The coupling parts 61 b and 81 b are connected to the web 202 in arotationally fixed but detachable manner. If the couplings are notengaged, the crosspiece and the ring gear can rotate freely.

The coupling components are numbered in the sequence in which theyfollow each other on axis 6 a from the drive side in steps of ten, eachstarting with 20 to 90; FIG. 1a -2 b.

Clutch 20 switches the block rotation of the first gear unit G1 byconnecting or disconnecting the sun gear 9 c to the drive 4 a and thusto the web 8.

Clutch 20 shifts the gears 1, 4, 7. The travel stop 21 a engages thecounterpart 21 b, which is firmly connected to the sun, and prevents thesun gear from overrunning the web 8.

The other components are a return spring 22, which is supported in theactuator 4 a, a slide 23 a with a cylindrical pin 23 b, which iscontrolled by the groove 24.

The clutch 30 shifts the second gear of the first transmission byshifting the sun wheel 9 c is set fixed to the axis. The clutch 30shifts the gears 2, 5, 7 of the overall transmission.

Clutch 40 shifts the third gear of the first transmission by setting thesun gear 9 d to fixed position. Clutch 40 shifts gears 3, 6, 9 of theoverall transmission.

The sun gear 9 d is designed as a ring with internal gearing in whichthe clutch part 41 a (FIG. 9) designed as an anti-running clutch isguided in a rotationally fixed but axially displaceable manner. Thecounterpart 41 b (FIG. 10) has internal teeth that connect it to theshaft 6 a in a rotationally secure manner. The retaining rings 44 a and44 b prevent axial movement. A slider 43 a with cylindrical pin 43 b iscontrolled by the groove 44. The spring 42 is supported by a circlip 42a inserted in the axle 6 a.

The clutches 50, 70, 80 shift the first gear of the second gear G2 byconnecting clutch parts 51 a (with cylindrical pin 51 c) and 51 bnon-rotatably connecting the ring gear with the driver 4 b, clutch parts81 b and 81 a (with cylindrical pin 81 c) non-rotatably connecting theweb with the output sleeve 4 d, and clutch parts 71 a (FIG. 11) and 71 bsetting the sun gear 204 axially fixed.

If this configuration is shifted, the gears 1, 2 and 3 are obtainedtogether with the first gear.

The other components of the clutch 50 are a return spring 52 supportedon the side wall 7 b, a groove ring 53 with inner collar, a slider 55 awith a cylindrical pin 55 b controlled by the groove 54.

The clutch 70 is designed as a bidirectional clutch, since the directionof the torque is reversed when the input is switched from the ring gearto the bar. The other components of the clutch 70 are, a return spring72, which is supported on the outer collar of the slider 63 a, a slider73 a with a cylindrical pin 73 b, which is controlled by the groove 74.

The other components of the clutch 80 are a return spring 82, which issupported on the outer collar of the slider 94 a, and a slider 83 a witha cylindrical pin 83 b, which is guided in the groove 84.

The couplings 50 and 90, FIG. 2a , connect the input sleeve 4 b with theoutput sleeve 4 c via the ring gear 200, thus bridging the gear unit.The same result is obtained by engaging clutches 60 and 80, with the webtransmitting the torque. In both cases, the sun can remain fixed to theaxle, but this is only used for shifting purposes for a short time dueto the loss of power.

The following options are available for block rotation, whereby the sunwheel can always rotate freely by opening clutch 70.

Variant 1: Clutches 50, 60, 90 closed, clutch 70 open.Variant 2: Clutches 50, 80, 90 closed, clutch 70 open.Variant 3: Clutches 50, 60, 80, 90 closed, clutch 70 open.Variant 4: Couplings 60, 80, 90 closed, coupling 70 open.Variant 5: Couplings 50, 60, 90 closed, coupling 70 open.Variant 4: Couplings 50, 60, 80 closed, coupling 70 open.

Some of the possibilities are used as examples in the suggestedswitching sequence. In this configuration the second gear of the secondtransmission is realized. Gears 4, 5 and 6 can then be shifted.

Clutches 60, 70, and 90 shift the third gear of the second transmissionby connecting clutch parts 61 a and 61 b to the web 202 with the driver4 b in a rotationally fixed manner, clutch parts 91 b and 91 a (withcylindrical pin 91 c) to connect the ring gear 200 with the outputsleeve 4 d in a rotationally fixed manner, and clutch parts 71 a and 71b to set the sun gear 204 in an axially fixed manner.

In this configuration, the third gear of the second gear is realized.Gears 7, 8 and 9 can then be shifted.

The solution of the task underlying the invention allows the use ofneedle bearings for the bearing of all planets. Due to the high loadcapacity of the manual transmission, ratios of up to i=1 can be usedbetween bottom bracket chainring and hub drive pinion. The exclusive useof axial clutches allows particularly good opening even under load. Inmost shifting operations, the clutches change to freewheel mode beforethey are opened, so that they can be opened without load. The number ofdifferent production parts is noticeably smaller than with the state ofthe art. The same applies to the total mass of the manual transmission.Due to the transmission structure, only a few gears are in mesh at anyone time, which ensures high efficiency in the individual gears.

FIG. 1a, 1b : Gearbox EG FIG. 2a, 2b : Gearbox NSG

FIG. 3A, 3 b: Switching sequence table with associated legend

FIG. 4: Total transmission in first gear

FIG. 5: Total transmission in 5th gear

FIG. 6: Total transmission in 9th gear

FIG. 7: Perspective view of the drum

FIG. 8: A side view of the shift drum

FIG. 9: A perspective view of coupling half 41 a

FIG. 10: A perspective view of the coupling half 41 b

FIG. 11: A perspective view of the clutch 70 with the sun wheel 204

1. Multi-speed transmission mounted on a central axis with a hollowshaft on the input side and a hub sleeve on the output side and with twocoaxially interposed or planetary gearboxes (EG, NSG) arranged coaxiallytherebetween or respectively therein, each with sun gear wheels,planetary gear wheels and hollow gear wheels connected to one anothervia associated webs, which can each be switched into a block revolutionand several stepped conversion modes, wherein the first planetarygearbox (EG) is driven at its web and can be switched into a blockrevolution and several stepped conversion modes by means of threeclutches can optionally be shifted into the block revolution or intosaid transmission modes by axle fixing of individual sun gears, and thesecond planetary gearbox (NSG) can be shifted into said transmissionmodes by means of further couplings by axle fixing of the sun gear withmutual shifting of the input and output between its ring gear and itsring gear respectively, and wherein the sun gear can be shifted into aslow and a fast mode as well as into the direct gear when the input andoutput are switched alternately between its ring gear and its web. 2.Multi-speed transmission according to claim 1, further comprisingaxially displaceable control slides, clutch rings extending radially andspring loaded in a coupling manner to be disengaged according to arespective gear step, wherein the control slides, guided in helicalcircular grooves in a coaxial shift drum and in radial slots of a hollowaxis (6 a) surrounding them and oriented parallel to the axis ofrotation of the transmission, each actuate the associated clutch ringsin a shifting manner.
 3. Multi-speed transmission according to claim 2,wherein the grooves are configured such a that a nine-speed transmissionwith 40 degrees angular rotation per gear in total within only onerevolution of the shift drum, can be set, whereby at least oneintermediate shift combination is passed through.
 4. Multi-speedtransmission according to claim 3, wherein in four of the nine gearsonly one planetary set is activated and in four others of the nine gearsboth planetary sets are activated and in addition one of the gears, thedirect gear, operates without rolling motion of the gears. 5.Multi-speed transmission according to claim 1, wherein the firsttransmission provides the speed conversions to i_1=1, i_2 approximately0.81 and i_3 approximately 0.65 and the second transmission (NSG)provides the speed conversions to i_4 approximately 1.91, i_5=1 and i_6approximately 0.52 such that there are largely constant gear steps ofapprox. 24% with a transmission ratio range of approximately 560%. 6.Multi-speed transmission according to claim 2, wherein the controlslides are each guided by a cylindrical pin engaging in the associatedgroove and the radial slot oriented parallel to the axis of rotation ofthe transmission.
 7. Multi-speed transmission according to the genericterm of claim 1, wherein the first planetary gearbox (EG) is a three- orfour-stage planetary gearbox, which is driven at its web and is switchedin each case by means of a clutch by fixing an associated sun wheel intoone of the said transmission modes, whereby a twelve- or fifteen-speedgearbox is formed in total.
 8. Multi-speed transmission according toclaim 2, wherein the clutch rings have a radial profiling which effectsthe power transmission.
 9. Multi-speed transmission according to claim8, wherein the radial profiling of the clutch rings, which are activatedin only one direction of rotation, are asymmetrically profiled such thatthe pressure-loaded flanks are steeper than the flanks on the rear sidethereof.
 10. Multi-speed transmission according to claim 1, wherein thesun, planetary and ring gears, which are activated in only one directionof rotation, and have an asymmetrical profiling of the wheel teeth suchthat the flanks loaded by tensile force have a larger tooth engagementangle than the flanks at the rear thereof.
 11. Multi-speed transmissionaccording to claim 1, wherein the driver is equipped with a toothed beltor chain pinion which has a has a number of teeth which correspondsapproximately to that of a correlated crank output blade. 12.Multi-speed transmission according to claim 11, wherein the toothed beltpinion has 34 teeth and the crank output blade has 42 teeth. 13.Multi-speed transmission according to claim 8, wherein the sun gear isaxially fixed in both directions of rotation by only one clutch. 14.Multi-speed transmission according to claim 8, wherein the clutch partsto be coupled in both directions of rotation have a symmetrical profileand can thus set the sun gear axially fixed in both directions ofrotation.
 15. According to claim 2, wherein the structures of thehelical grooves in the shift drum as well as the slots in the hollowaxis, which serve for shifting all gears, are arranged on a maximum of180 degrees of their circumferences and copies thereof are arranged ontheir other circumferential halves, in each of which a furthercylindrical pin engages, which in each case actuates the associatedcoupling ring in parallel on the other side.